U.S. patent application number 17/235167 was filed with the patent office on 2021-10-28 for valve poppets and valve seats for high-speed reciprocating compressor capacity unloaders.
The applicant listed for this patent is ACI SERVICES, INC.. Invention is credited to David W. Legg, W. Norman Shade.
Application Number | 20210332902 17/235167 |
Document ID | / |
Family ID | 1000005598165 |
Filed Date | 2021-10-28 |
United States Patent
Application |
20210332902 |
Kind Code |
A1 |
Shade; W. Norman ; et
al. |
October 28, 2021 |
VALVE POPPETS AND VALVE SEATS FOR HIGH-SPEED RECIPROCATING
COMPRESSOR CAPACITY UNLOADERS
Abstract
An improved unloader valve assembly for a high-speed
reciprocating compressor includes improved poppets and valve seat
designs. The improved poppets and valve seats are useful for
reducing or eliminating poppet leakage and high impact stresses on
the sealing surfaces of the poppets, and are especially intended
for use with high-speed reciprocating compressors operating at 1000
rpm and higher. A control chamber spacer plate for creating a
control pressure chamber within the unloader valve assembly, and a
valve seat cushioning plate for reducing stress on the head ends of
the poppets are also described.
Inventors: |
Shade; W. Norman; (Port
Clinton, OH) ; Legg; David W.; (Marietta,
OH) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
ACI SERVICES, INC. |
Cambridge |
OH |
US |
|
|
Family ID: |
1000005598165 |
Appl. No.: |
17/235167 |
Filed: |
April 20, 2021 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
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63014293 |
Apr 23, 2020 |
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Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F16K 5/0464 20130101;
F16K 17/10 20130101; F04B 39/1013 20130101 |
International
Class: |
F16K 17/10 20060101
F16K017/10; F04B 39/10 20060101 F04B039/10; F16K 5/04 20060101
F16K005/04 |
Claims
1. An unloader valve assembly for a reciprocating compressor,
comprising: a) a clearance pocket having a volume; b) a support
dome including a plurality of support dome ports, wherein each of
the plurality of support dome ports communicates with the clearance
pocket; c) a valve guard including: i) a plurality of poppet guide
recesses; ii) a plurality of poppets, wherein each of the plurality
of recesses houses one of the plurality of poppets, each of the
plurality of poppets having an upper sealing surface and a lower
sealing surface; iii) a plurality of valve guard seating surfaces
for contacting the upper sealing surfaces of the plurality of
poppets, wherein each of the plurality of valve guard seating
surfaces contacts an upper sealing surface of one of the plurality
of poppets; iv) a spacer plate portion comprising a plurality of
posts for supporting the valve guard against the support dome,
wherein the spacer plate and the support dome form a control
pressure chamber, wherein each of the plurality of posts include a
central port which communicates with one of the plurality of
support dome ports; v) a plurality of control pressure ports, each
of the plurality of control pressure ports located in the spacer
plate portion between the plurality of posts, wherein each of the
plurality of control pressure ports communicates with the control
pressure chamber and with one of the plurality of poppet guide
recesses; and d) a valve seat including: i) a valve seat cushioning
plate for reducing the magnitude of impact stresses acting on the
head end of the poppet, the valve seat cushioning plate including a
plurality of valve seat seating surfaces for contacting the lower
sealing surfaces of the plurality of poppets, wherein each of the
plurality of valve seat seating surfaces contacts a lower sealing
surface of one of the plurality of poppets; and ii) a plurality of
valve seat ports, wherein each of the plurality of valve seat ports
communicates with a cylinder of a reciprocating compressor.
2. The unloader valve assembly of claim 1, wherein each of the
plurality of posts of the spacer plate has a truncated triangle
shape for augmenting the support area provided by the posts.
3. The unloader valve assembly of claim 1, the valve seat
cushioning plate further comprising a cushioning element, wherein
the cushioning element is selected from the group consisting of a
cushioning pad, an O-ring, one or more springs, or a combination of
thereof.
4. The unloader valve assembly of claim 1, wherein the diameter of
each of the plurality of valve seat ports is between 0.380 inches
and 0.410 inches, and more preferably 0.400 inches, and wherein the
diameter of each of the plurality of control pressure ports is
between 0.380 inches and 0.410 inches, and more preferably 0.400
inches.
5. The unloader valve assembly of claim 4, wherein contact between
each of the plurality of lower sealing surfaces and valve seat
seating surface creates a circular contact line on the lower
sealing surface of each of the plurality of poppets, wherein the
mean diameter of the circular contact line of the lower sealing
surface is between 0.020 inches and 0.045 inches larger than the
valve seat port diameter, more preferably between 0.030 inches and
0.044 inches larger than the valve seat port diameter, and most
preferably between 0.038 inches and 0.042 inches larger than the
valve seat port diameter.
6. The unloader valve assembly of claim 5, wherein the traveling
distance of each of the plurality of poppets multiplied by the
circumference of the inner edge of the (original) circular contact
line of the lower sealing surface of each of the poppets is between
97% and 103% of the valve seat port area.
7. The unloader valve assembly of claim 4, wherein contact between
each of the plurality of upper sealing surfaces and valve guard
seating surface creates a circular contact line on the upper
sealing surface of each of the plurality of poppets, wherein the
mean diameter of the circular contact line of the upper sealing
surface is between 0.020 inches and 0.045 inches larger than the
control pressure port diameter, more preferably between 0.030
inches and 0.044 inches larger than the control pressure port
diameter, and most preferably between 0.038 inches and 0.042 inches
larger than the control pressure port diameter.
8. The unloader valve assembly of claim 7, wherein the traveling
distance of each of the plurality of poppets multiplied by the
circumference of the inner diameter of the circular contact line of
the upper sealing surface of each of the poppets is between 97% and
103% of the control pressure port area.
9. The unloader valve assembly of claim 1, wherein the angle of
each of the plurality of valve seat seating surfaces is between 45
degrees and 35 degrees, preferably between 42.5 degrees and 37.5
degrees, and most preferably 40 degrees, and wherein the angle of
each of the plurality of valve guard seating surfaces is between 45
degrees and 35 degrees, preferably between 42.5 degrees and 37.5
degrees, and most preferably 40 degrees.
10. The unloader valve assembly of claim 9, wherein the angle of
each of the plurality of lower sealing surfaces is between 1.0
degree and 5.0 degrees larger than of the angle of each of the
plurality of valve seat seating surfaces, and preferably 3.0
degrees larger than the angle of each of the plurality of valve
seat seating surfaces, and wherein the angle each of the plurality
of upper sealing surfaces is between 1.0 degree and 5.0 degrees
larger than the angle of each of the plurality of valve guard
seating surfaces, and preferably 3.0 degrees larger than the angle
of each of the plurality of valve guard seating surfaces.
11. The unloader valve assembly of claim 1, wherein each of the
plurality of poppets is a pressure breaker poppet, each pressure
breaker poppet comprising: a) a head end having a first diameter,
the head end including: i) the upper sealing surface for contacting
one of the plurality of valve guard seating surfaces, wherein
contact between the upper sealing surface and the valve guard
seating surface creates a contact line on the upper sealing
surface; ii) the lower sealing surface for contacting one of the
plurality of valve seat seating surfaces, wherein contact between
the lower sealing surface and the valve seat seating surface
creates a contact line on the lower sealing surface; b) a stem end
having a second diameter which is smaller than the first diameter
of the head end, the stem end including a plurality of steps
radially protruding from the second diameter, wherein the plurality
of steps are sized to fit inside one of the plurality of poppet
guide recesses to reduce leakage around the poppet; and c) a
sealing guide stem having a third diameter which is smaller than
the second diameter of the stem end, wherein the sealing guide stem
is sized to fit within one of the plurality of control pressure
ports in the spacer plate portion to reduce leakage around the
poppet.
12. The unloader valve assembly of claim 11, wherein each of the
plurality of steps are radial protrusions selected from the group
consisting of square edges, sharp-edged labyrinth teeth, slanted
protrusions with labyrinth teeth, or a combination thereof.
13. The unloader valve assembly of claim 11, wherein the poppet
further includes an internal hollow volume for limiting the poppet
mass and the impact stresses on the sealing surfaces.
14. The unloader valve assembly of claim 11, wherein the first
diameter of the head end of the poppet is between 0.510 inches and
0.490 inches, more preferably between 0.502 inches and 0.498
inches, and most preferably between 0.501 inches and 0.499
inches.
15. The unloader valve assembly of claim 1, wherein each of the
plurality of poppets is a pressure breaker poppet, each pressure
breaker poppet comprising: a) a head end having a first diameter,
the head end including: i) the upper sealing surface for contacting
one of the plurality of valve guard seating surfaces, wherein
contact between the upper sealing surface and the valve guard
seating surface creates a contact line on the upper sealing
surface; ii) the lower sealing surface for contacting one of the
plurality of valve seat seating surfaces, wherein contact between
the lower sealing surface and the valve seat seating surface
creates a contact line on the lower sealing surface; and b) a stem
end having a second diameter which is smaller than the first
diameter of the head end, the stem end including a plurality of
steps radially protruding from the second diameter, wherein the
plurality of steps are sized to fit inside one of the plurality of
poppet guide recesses to reduce leakage around the poppet.
16. The unloader valve assembly of claim 15, wherein each of the
plurality of steps of the poppet are radial protrusions selected
from the group consisting of square edges, sharp-edged labyrinth
teeth, slanted protrusions with labyrinth teeth, or a combination
thereof.
17. The unloader valve assembly of claim 15, wherein the poppet
further includes an internal hollow volume for limiting the poppet
mass and the impact stresses on the sealing surfaces.
18. The unloader valve assembly of claim 15, wherein the first
diameter of the head end of the poppet is between 0.510 inches and
0.490 inches, more preferably between 0.502 inches and 0.498
inches, and most preferably between 0.501 inches and 0.499
inches.
19. The unloader valve assembly of claim 1, wherein each of the
plurality of poppets is an impact tolerant self-sealing poppet,
each impact tolerant self-sealing poppet comprising: a) a head end
having an outer diameter, the head end including a plurality of
steps radially protruding from the outer diameter of the head end;
and b) a stem end having an outer diameter that is substantially
the same as the outer diameter of the head end, the stem end
including a plurality of steps radially protruding from the outer
diameter of the stem end.
20. The unloader valve assembly of claim 19, wherein each of the
plurality of steps are radial protrusions selected from the group
consisting of square edges, sharp-edged labyrinth teeth, slanted
protrusions with labyrinth teeth, or a combination thereof.
21. The unloader valve assembly of claim 19, wherein the poppet
further includes an internal hollow volume for limiting the poppet
mass and the impact stresses on the sealing surfaces.
22. The unloader valve assembly of claim 19, wherein the outer
diameter of the head end and of the stem end of the poppet is
between 0.510 inches and 0.490 inches, more preferably between
0.502 inches and 0.498 inches, and most preferably between 0.501
inches and 0.499 inches.
23. The unloader valve assembly of claim 1, wherein each of the
plurality of poppets is an impact tolerant self-sealing poppet,
each impact tolerant self-sealing poppet comprising: a) a head end
having an outer diameter, the head end comprising a head end piece
including a central hollow portion having an inner diameter; b) a
stem end having an outer diameter that is substantially the same as
the outer diameter of the head end, the stem end comprising a stem
end piece including a central hollow portion having an inner
diameter that is substantially the same as the inner diameter of
the head end piece; and c) a central core positioned between the
head end piece and the stem end piece, the central core comprising
a hollow portion, a plurality of steps radially protruding from the
outer diameter of the central core, and bulbous-shaped ends for
connecting to the head end piece and the stem end piece, wherein
the bulbous-shaped ends have a larger outer diameter than the inner
diameters of the head end piece and the stem end piece.
24. The unloader valve assembly of claim 1, wherein each of the
plurality of poppets is a diaphragm seal poppet, each diaphragm
seal poppet comprising: a) a head end having an outer diameter; b)
a stem end having an outer diameter that is substantially the same
as the outer diameter of the head end; and c) a flexible diaphragm
seal projecting from the stem end, the diaphragm seal comprising:
i) a bulbous outer diameter clamped into a recess of the valve
guard; ii) a seating recess for contacting the valve guard seating
surface; and iii) at least one strain relief loop positioned
between the bulbous outer diameter and the seating recess.
25-37. (canceled)
Description
CROSS-REFERENCE TO RELATED APPLICATION
[0001] The present application claims the benefit of U.S.
Provisional Application Ser. No. 63/014,293 filed Apr. 23, 2020,
the disclosure of which is hereby incorporated by reference in its
entirety.
FIELD OF THE INVENTION
[0002] The present invention relates generally to capacity
unloaders for high-speed reciprocating compressors, and in
particular to improved poppets and valve seats for the automatic
variation of fixed clearance volumes with high-speed reciprocating
compressors.
BACKGROUND OF THE INVENTION
[0003] Reciprocating compressors are positive displacement machines
wherein a reciprocating piston moves back and forth within a fixed
cylindrical volume. Specifically, most known reciprocating
compressors typically include a cylinder, a piston reciprocally
mounted in the cylinder, a rotatable crankshaft driven by an engine
or electric motor and connected to the piston, and a suction valve
assembly and a discharge valve assembly for selectively
communicating suction and discharge lines with the compressor
cylinder. Reciprocating compressors are commonly used for a wide
range of applications such as the pressurization and transport of
natural gas and mixtures of gases through systems used for
industrial and commercial processes.
[0004] In natural gas applications, transmission through pipelines
is commonly accomplished with large reciprocating compressors
driven by internal combustion engines at pumping stations located
along the pipeline routes. Reciprocating compressors can also be
driven by electric motors and other drivers, and they are commonly
employed in refineries and chemical process plants to pressurize
and move natural gas, hydrogen and many other gases throughout the
process facility. Other examples include, but are not limited to,
industrial air compression, process refrigeration, and vapor
recovery from storage tanks, operating equipment and other
processes.
[0005] Whether driven by internal combustion engines, electric
motors or other means, a reciprocating compressor's capacity is
directly related to the fixed geometry built into its compressor
cylinder(s). Defined as the total flow rate or output of a
compressor, compressor capacity is a function of cylinder
displacement and the internal clearance volume remaining in the
cylinder when the piston is at the end of its stroke. Cylinder
displacement equals the area of the piston end face multiplied by
the length of the stroke of the piston.
[0006] The extent to which a reciprocating compressor is loaded can
have a significant effect on its energy consumption and thus
compressor operating efficiency, cost-effectiveness and
environmental impact. In the fully-loaded condition, the maximum
output of the compressor is achieved, with a resultant full load on
the compressor's driver. However, gas flow and pressure
requirements can vary considerably, typically depending upon
upstream supply conditions as well as downstream demand
conditions.
[0007] Controlling compressor flow is often accomplished by
partially "unloading" a compressor, whereby each compressor stroke
produces a reduced gas flow as compared to fully-loaded operation.
Reduced gas flow generally corresponds to reduced work performed by
the compressor, such that fuel savings and greater efficiency can
be achieved. Although compressor output can be varied by changing
the speed of the driving engine, this approach can be impractical
because the engines are typically designed to operate at constant
speeds for maximum fuel efficiency and minimum emissions. Thus,
compressor capacity and flow rate control must normally be
accomplished using other means, such as by changing the internal
clearance volume of the system.
[0008] A compressor can be partially unloaded and its output
reduced by increasing the clearance volume. Clearance pockets or
bottles connected to the compressor cylinder via an unloader valve
are often provided for this purpose, for example, either via a
fixed volume clearance pocket or a variable volume clearance
pocket, either of which are typically located at the outer head of
the compressor cylinder. Adding clearance volume reduces the
compressor capacity, and removing clearance volume increases the
compressor capacity. Clearance pockets (fixed and variable) are
commonly referred to as "unloaders" because they can reduce the
capacity and therefore "unload" the compressor; manipulation of
clearance pockets can also "load" the compressor by removing
clearance volume and thus increasing the capacity.
[0009] The clearance volume provided by fixed means must be either
fully added or fully removed prior to or during operation of the
compressor, and cannot be used in a variable manner when the
compressor is in operation. In contrast, the volume of a variable
clearance pocket can be changed while the compressor is in
continuous operation, and is one of the most effective means of
changing the compressor capacity and the required power in real
time. A variable clearance pocket can be positioned at an infinite
number of positions or steps within the range of clearance volumes
it is designed to provide. Such devices have been in use throughout
the compressor industry for many years.
[0010] For example, U.S. Pat. Nos. 6,361,288 and 6,607,366 both to
Sperry disclose variable clearance volume systems for reciprocating
compressors in which an unloader valve assembly (including a valve
seat structure, a valve guard and multiple poppet valve members) is
provided to allow selective communication between the compressor
cylinder and a separate clearance volume. The opening and closing
of the unloader valve assembly is controlled by manipulating a
control pressure regulator connected in series with a pressure
source. When the pressure in the cylinder acting on the heads of
the poppet valve members exceeds the control pressure acting on the
stems, the poppet valve members open, partially unloading the
compressor.
[0011] The prior art Sperry variable clearance systems can
pneumatically load and unload a reciprocating compressor in a
smooth, stepless manner with each revolution of the crankshaft.
This is accomplished by using a controlled pressure to hold the
unloader poppets closed until the compressor piston reaches the
desired position of its cycle. By adjusting the set point of a
pressure regulator, the effective use of any shape and size of
clearance cavity can be smoothly varied from having no effect to
having full effect.
[0012] The Sperry variable clearance system discussed above was
originally developed and used for reciprocating compressors
operating generally at 200 to about 400 rpm. Subsequent
improvements to the system, including those disclosed by Sanford
(U.S. Pat. No. 8,070,461), have resulted in limited use of the
concept on compressors up at about 900 to 1000 rpm. However,
reciprocating compressors have evolved to higher and higher speeds.
In particular, large reciprocating compressors used in natural gas
production, gathering and pressure boosting at or near oil and gas
well-heads now commonly operate at speeds of higher than 1000 rpm,
most commonly at speeds of 1200 and 1400 rpm, and in many cases, at
speeds of 1800 rpm or higher.
[0013] At these higher speeds, prior art systems generally
encounter significant operating limitations that render them
ineffective and unreliable. Specifically, response time and
reliability of the poppets can limit the application of the Sperry
technology at speeds above about 1000 rpm. Leakage of gas past the
poppet head seats in the closed position and the poppet stem seats
in the open position affects the control pressure, limiting and
reducing the effective unloading of the compressor compared to the
desired set point. In addition, higher operating speeds increase
the impact forces and therefore the impact stresses on the poppets
and seats, which can lead to failure of compressor components and
reduced reliability and safety.
[0014] Reciprocating Compressor Compression Cycle--A quick
explanation of a few basic thermodynamic principles is necessary to
understand the science of reciprocating compressors. Referring now
to FIGS. 1A-1F depicting various stages of a compressor cylinder 1
as a reciprocating piston 2 moves back and forth within it. The
system also includes a suction valve 3, a discharge valve 4, and a
fixed volume clearance pocket 5. The internal volume of a fixed
volume clearance pocket 5 is connected to the internal cylinder
volume 6 by a port 7. The port is opened or closed by a plug 8,
which is connected to an actuator 9 which is manipulated manually
or automatically by various means.
[0015] Compression occurs within the cylinder as a four-part cycle
that occurs with each advance and retreat of the piston (two
strokes per cycle). The four parts of the cycle are compression,
discharge, expansion and intake. See FIG. 2, which shows the
operation of the cylinder 1 of FIG. 1 graphically with pressure vs.
volume plotted in what is known as a P-V diagram. Specifically, the
outer, "fully loaded" pressure-volume envelope in FIG. 2 depicts
the operation of the cylinder from point 1 to point 2 to point 3 to
point 4 and back to point 1 with one complete revolution of the
compressor crankshaft, and with port 7 closed off by plug 8 so that
the added clearance pocket volume 5 is not in communication with
the cylinder internal volume 6. The inner, "unloaded"
pressure-volume envelope in FIG. 2 depicts the operation of the
cylinder from point 1 to point A to point 3 to point B and back to
point 1 with one complete revolution of the compressor crankshaft
with port 7 opened, so that the added clearance pocket volume 5 is
in communication with the cylinder internal volume 6.
[0016] At the conclusion of a prior cycle, the piston 2 in FIG. 1A
is at the limit of its travel away from the closed end of the
cylinder 1, commonly referred to as inner dead center and
represented by point 1 in FIG. 2. The cylinder's internal volume 6
is filled with process gas at suction conditions (as shown in FIG.
2 at 750 psig), and the suction 3 and discharge 4 valves are
closed. As the piston 2 begins to advance toward the closed end of
the cylinder, the volume 6 inside the cylinder is reduced and the
pressure is increased. The increased cylinder pressure causes the
suction valve 3 to close and, with port 7 closed by plug 8, the
internal volume 6 decreases along the pressure-volume path from
point 1 to point 2. This is referred to as the "compression
stroke." The pressure inside the cylinder increases until the
pressure within the cylinder exceeds the discharge pressure (e.g.,
1250 psig in FIG. 2) and reaches the pressure required to open the
discharge valve 4.
[0017] At point 2 in FIG. 2 the increased pressure forces the
discharge valve 4 open and gas flows out of the cylinder. Cylinder
pressure decreases slightly for the remainder of the advancing
stroke as volume continues to decrease for the discharge portion of
the cycle. The cycle is now at operating point 3 in FIG. 2. The
piston comes to a momentary stop at the most advanced position in
its travel, as shown in FIG. 1B, which is commonly referred to as
outer dead center, before reversing direction. The pressure-volume
path from point 2 to point 3 in FIG. 2 is referred to as the
"discharge stroke."
[0018] Note that some minimal volume remains in the cylinder in
FIG. 1B, known as the clearance volume. It is the space remaining
within the cylinder when the piston is at point 3, after the
compressed gas is discharged from the cylinder. Some minimum
clearance volume is necessary to prevent piston/head contact, and
the efficient manipulation of the clearance volume is a major
parameter of compressor performance.
[0019] As the piston 2 begins its return stroke away from the
closed end of the cylinder as shown in FIG. 1C, the discharge valve
4 closes, the volume 6 expands and, with port 7 closed by plug 8,
the pressure decreases along the path from point 3 to point 4 in
FIG. 2. The pressure-volume path from point 3 to point 4 is
referred to as the "expansion stroke." The gas which remained in
this space re-expands to slightly below suction pressure (e.g. 750
psig in FIG. 2). Gas expansion within the cylinder is facilitated
by the closing of the discharge valve 4 and the retreat of the
piston 2. At point 4, the low internal cylinder pressure causes the
suction valve 3 to open and fresh gas is admitted into the
cylinder, as shown in FIG. 1D, until the piston 2 reaches the limit
of its travel away from the closed end of the cylinder 1. The
pressure-volume path from point 4 to point 1 in FIG. 2 is referred
to as the "suction stroke," Once again, pressure remains relatively
constant as the volume is changed. This marks the return to point
1. Comprehending this cycle is fundamental to diagnosing compressor
problems, and to understanding compressor efficiency, power
requirements, valve operation, etc.
[0020] Reciprocating Compressor Unloading with a Fixed Volume
Clearance Pocket--The connection of a fixed volume clearance pocket
is a commonly known and applied method of reducing the flow, or
capacity, and the load, or power requirement, of a reciprocating
compressor. This is demonstrated in FIG. 1E, FIG. 1F and FIG. 2. As
shown in FIG. 1E with the plug 8 moved away from open port 7 such
that the added clearance pocket volume 5 is in communication with
the cylinder internal volume 6, as the piston 2 moves toward the
closed end of the cylinder, the pressure is increased in both the
internal cylinder volume 6 and in the pocket volume 5. Since the
combined volume of 6 and 5 is larger than the internal cylinder
volume 6 alone, the pressure-volume path from point 1 to point A in
FIG. 2 shows that more time and more piston travel is required to
compress the larger combined volume to increase the pressure to a
level that is sufficient to open the discharge valve.
[0021] Similarly, as shown in FIG. 1F with the plug 8 moved away
from open port 7 such that the added clearance pocket volume 5 is
in communication with the cylinder internal volume 6, as the piston
2 travels away from the closed end of the cylinder, the pressure
decreases in both the internal cylinder volume 6 and in the pocket
volume 5. Since the combined volume of 6 and 5 is larger than the
internal cylinder volume 6 alone, the pressure-volume path from
point 3 to point B in FIG. 2 shows that more time and more piston
travel is required to expand the larger combined volume to decrease
the pressure to a level sufficient to open the suction valve 3.
[0022] As an example of how adding clearance volume affects
reciprocating compressor capacity and required power, the
compressor represented by the P-V diagram in FIG. 2 has a cylinder
with a 7 in. diameter bore and a 6.5 in. stroke, operating at 1000
rpm to deliver gas from a 750 psig suction pressure to a 1250 psig
discharge pressure. In the fully-loaded state at these operating
conditions, the compressor produces a fully loaded capacity of
10.18 million standard cubic feet per day (MMCFD) and requires
289.7 horsepower (HP). Connecting a 250 in.sup.3 clearance volume
pocket to this cylinder, at the same operating speed and pressures,
reduces the flow to 4.56 MMSCFD (44.8% of full-load capacity) and
requires 118.2 HP (40.8% of rated power). Notably on a percentage
basis the power is reduced more than the capacity, since the lower
flow rate generates a lower pressure resistance, and therefore
lower parasitic losses through the suction and the discharge
valves. Accordingly, the addition of clearance volume is a
preferred method of reducing the reciprocating compressor capacity
and the power required for compression.
[0023] Variable Volume Clearance Pockets--As demonstrated by the
foregoing discussion of FIGS. 1 and 2, the clearance volume
addition can be a fixed amount that causes a significant but fixed
step change in capacity and required power. When considering
variable operating requirements, it is desirable to be able to
change the compressor's capacity and required power in a
continuously variable or "stepless" manner, so as to operate the
compressor at maximum efficiency as operating conditions and flow
requirements change. Since it is not practical to add a plurality
of small fixed volume clearance pockets to a single cylinder in a
manner that could effectively provide very small steps of clearance
volume adjustment, a means of adjusting the added clearance volume
in very small steps and in a continuously variable or "stepless"
manner over a range is needed.
[0024] Manual means of such adjustment have been commonly applied
for many decades. However, manual adjustment requires a manual
intervention and effort, which is not practical for compressors
that are unattended, that have frequently changing operating
requirements, and/or that are expected to operate automatically and
continuously with frequent and/or sudden changes in operating
conditions. In fact, a large majority of all reciprocating
compressors are operated unattended.
[0025] Various means of automatic clearance adjustments have been
developed, for example U.S. Pat. No. 8,430,646 assigned to ACI
Services, Inc., which is incorporated herein by reference in its
entirety. However, such automatic means of clearance adjustment
require external motive power, such as pressurized air, pressurized
hydraulic fluid and/or electricity for operation, which can add
significant cost and complexity to the operation. In addition, in
remote applications, such as where oil and gas production
operations are commonly located and which require large numbers of
reciprocating compressors, electrical power is limited or
non-existent. Therefore, without electrical power to drive the
automatic clearance actuation directly, or to drive an air
compressor or a hydraulic pump to provide pressurized air or fluid,
respectively, for driving an actuator for the automatic clearance
adjustment, such automatic clearance devices are not practical. In
addition, the actuators or drivers for such automatic clearance
adjustment means tend to be large and heavy. The only accessible
location for mounting them is outboard of the cylinder, which
extends its length in the direction of cylinder piston travel. The
increased mass and extension of the length of the cylinder assembly
amplify the normal mechanical vibration that occurs as a result of
the reciprocating inertia and pressure forces acting on the
compressor cylinder, and they reduce the cylinder's mechanical
natural frequency, which can make it more likely to experience
unsafe levels of mechanical vibration during operation.
[0026] Operation and Advantages of the Gas Operated Variable
Clearance Volume Unloader for Variable Capacity Control--The
present invention is intended for high-speed use (i.e. 1000 rpm and
greater) in unloader valve assemblies currently employed in prior
art variable clearance systems, specifically, as disclosed at
columns 4-6 and in FIGS. 1-5 of U.S. Pat. No. 6,607,366 to Sperry
and incorporated herein by reference, an unloader system for a
reciprocating compressor. Referring to FIGS. 3 and 4 of the present
disclosure, which illustrate a depiction of such a prior art
unloader valve assembly 10 for a reciprocating compressor. The
reciprocating compressor includes a cylinder C, a piston P
reciprocally mounted in the cylinder C having a volume 14, a
rotatable crankshaft connected to a piston rod R that is connected
to the piston P, a suction valve assembly S, and a discharge valve
assembly D for selectively communicating suction and discharge
lines with the compressor cylinder. The unloader valve assembly 10
includes a valve seat structure 11, a valve guard 12, and multiple
poppets 13 (a "poppet" as described herein is also referred to in
the prior art as a poppet valve member, a closing element for a
valve assembly, or a valve poppet) to allow selective communication
between the compressor cylinder volume 14 and a separate clearance
volume 15.
[0027] Opening and closing of the poppets 13 is controlled by
manipulating a control pressure (P.sub.c, P.sub.cont) acting
against the stem ends 18 of the poppets 13. The control pressure
(P.sub.c) is provided by means of a control pressure chamber 17,
which is connected to a pressure source higher than the cylinder
pressure (P.sub.cyl). For example, the control pressure chamber 17
can be connected in series to a pressure control valve or pressure
regulator 16, as is known in the art. The activating and
deactivating of the clearance pocket volume 15 is triggered by the
magnitude of the control pressure (P.sub.c) in the control pressure
chamber 17, which can be steplessly varied between a pressure less
than the suction pressure and a pressure higher than the discharge
pressure of the cylinder (P.sub.cyl). As can be seen in FIG. 3,
when the control pressure (P.sub.c) in the control pressure chamber
17 is higher than the pressure (P.sub.cyl) in the cylinder volume
14, the head ends (i.e. nearest the piston P, and also referred to
as the HE) of the poppets 13 are pressed against the valve seat 11,
such that ports 19 in the valve seat 11 are sealed off, preventing
gas in the compressor cylinder volume 14 from communicating with
the clearance pocket volume 15. However, as shown in FIG. 4, when
the pressure in the cylinder volume 14 acting on the heads of the
poppets 13 exceeds the control pressure 17 acting on the stems 18
of the poppets, the stem ends 18 of the poppets are pressed against
the valve guard 12 causing the ports 19 in the valve seat 11 to be
opened, such that the compressor cylinder volume 14 is placed in
communication with the separate clearance volume 15, partially
unloading the compressor system.
[0028] FIGS. 5A to 5F show the P-V diagrams of a typical prior art
reciprocating compressor cylinder having a 7 in. bore diameter and
a 6.5 in. stroke, operating at 1000 rpm to deliver gas from a
suction pressure of 750 psig to a discharge pressure of 1250 psig.
This is the same compressor and operating conditions as represented
in FIG. 2. However, in the case represented in FIGS. 5A-5F, a 250
in.sup.3 clearance pocket 15 is connected to the compressor
cylinder volume 14 utilizing the prior art variable clearance
system described above, to provide various control pressure levels
acting on the stem ends 18 of the poppets 13 shown in FIG. 3 and
FIG. 4. FIG. 5A shows the P-V relationship of the cylinder with the
control pressure P.sub.c set at 1350 psig, which is higher than the
maximum internal cylinder pressure when operating with a 1250 psig
discharge pressure. At this control pressure, the opposite, head
ends of the poppets 13 remain pressed against the valve seat 11 for
the entire cycle, preventing communication of the cylinder volume
14 with the clearance pocket volume 15. See FIG. 3. In this "fully
loaded" condition, the cylinder produces 100% of rated flow and
requires 100% of rated compression power.
[0029] FIG. 5B shows the P-V diagram of this same cylinder
operating at the same conditions, but with the control pressure Pc
set at 1113 psig. At cylinder pressures higher than the 1113 psig
control pressure, the stem ends 18 of the poppets 13 are pressed
against the guard seat 12, opening the ports 19 in the valve seat
11 such that the internal cylinder volume is in communication with
the clearance pocket volume, as seen in FIG. 4. When the cylinder
pressure is less than the control pressure of 1113 psig the system
resembles FIG. 3, with the heads of the poppets 13 pressed against
the valve seat 11 closing the ports 19 in the valve seat such that
the internal cylinder volume 14 does not communicate with the
clearance pocket volume 15. The shape of the expansion and
compression lines change when the valve seat ports 19 are open to
enable the cylinder internal volume 14 to communicate with the
clearance pocket volume 15. This has the effect of unloading the
cylinder to reduce the flow and power to 88.3% and 82.6%,
respectively (see FIG. 5B).
[0030] The control pressure Pc can be changed to any level between
higher than discharge pressure, up to the maximum design pressure
of the assembly, and lower than suction pressure. For illustration
purposes, FIG. 5B to FIG. 5F show the P-V characteristics of the
cylinder with the control pressure Pc set progressively lower, i.e.
to 1113, 1000, 903, 817 and 750 psig, respectively. These control
pressure settings progressively unload the cylinder to flows of
88.3%, 78.1%, 67.2%, 56.3% and 44.8%, respectively, of rated flow,
with the power reduced to 82.6%, 70.5%, 60.0%, 50.5% and 40.8%,
respectively, of rated power. With the control pressure Pc set at
suction pressure (750 psig), as represented in FIG. 5F, the
cylinder internal volume 14 is in communication with the clearance
pocket volume 15 for the entire cycle. The cylinder flow and power
unloading achieved with this prior art system in FIG. 5F is exactly
the same as the flow and unloading for the fixed volume clearance
pocket of the same volume in FIG. 2.
[0031] By comparing FIG. 5A with FIGS. 5B through 5F, which have
progressively lower control pressure settings, activating the
clearance volume 15 at a certain pressure in the control chamber 17
as represented in FIG. 4, it can be appreciated that the slope of
the compression stroke becomes smaller when the clearance volume
pocket 15 is communicating with the compressor cylinder volume 14.
As a result, the discharge volume decreases. To adjust the reduced
discharge volume on the expansion stroke, the clearance pocket 15
must be deactivated (i.e., no communication with the compressor
cylinder volume 14) at the same pressure level, such that the slope
of the expansion line becomes larger.
[0032] Accordingly, the prior art variable clearance system
discussed above enables operation of the compressor cylinder in a
fully loaded condition, a fully unloaded condition (dependent on
the volume of the fixed clearance pocket), and at any partially
unloaded condition between fully loaded and fully unloaded, as
determined by the control pressure setting. This variable clearance
system has several advantages over other automatic variable
clearance volume systems. First, it requires only the compressed
gas, supplied at a pressure slightly higher than the stage's
discharge pressure, for operation. It does not require a
high-voltage electrical supply, a pneumatic system, or a hydraulic
system, with their attendant secondary motive power and control
systems. Second, this variable clearance system is completely
sealed and results in no gas emissions, venting or leakage to the
atmosphere during operation. Therefore, in addition to being more
environmentally compliant, the prior art Sperry variable clearance
system is simpler and typically has much lower initial cost and
lower operating and maintenance costs than any other automatic
variable clearance system that has been conceived thus far.
[0033] Limitations of the prior art variable clearance system--The
prior art variable clearance systems discussed above were developed
and applied to compressors operating generally in the 200 to 400
rpm range. This is known by the present inventors who worked
directly with inventor Lauren D. Sperry for years. This fact is
further evidenced by the large, mushroom-head poppets that are
shown in FIG. 2 of U.S. Pat. Nos. 6,361,288 and 6,607,366 to
Sperry, and incorporated by reference herein in their entirety. The
head diameter (item 43 in FIG. 2 of the referenced patents) is
typically either 1.12 inches or 1.38 inches. Years of application
and operating experience has revealed to the present inventors
several limitations in these prior art variable clearance systems,
and these limitations are exacerbated when used with compressors
operating at speeds higher than 400 rpm.
[0034] Limitation 1. High-speed compressors are very prevalent in
most gas compression applications at the current time, with rated
operating speeds above 1000 rpm quite common, and speeds of 1200 to
1800 rpm most common. As the speed increases, the cycle time
required is proportionally shorter for the poppet valve members
(referred to herein as "poppets") to traverse the gap between the
seating surfaces in the valve guard and the valve seat. In
addition, the velocity, acceleration, and deceleration of the
poppets are much higher at speeds above 1000 rpm than with the
longer poppet opening and closing cycle times at lower speeds. The
impact stress on the poppet seating face, which is generally the
surface undergoing the highest stress in the poppet (see FIGS.
6A/B), is proportional to the poppet mass and to the square of the
impact velocity. High impact velocity and the resulting high
stresses on the poppet seating faces typically lead to premature
failures of the poppets, and therefore, of the unloading system.
FIG. 6A illustrates an example of 0.500 in. diameter poppets 13
from a prior art variable clearance system and two failed poppets
21. The head ends of the failed poppets 21 become cracked, worn
out, distorted, or otherwise broken within about two hours of
testing on a reciprocating compressor operating at 1200 rpm with a
control pressure of 1232 psig, a compressor cylinder suction
pressure of 835 psig and a discharge pressure of 1345 psig. The
accompanying stress map in FIG. 6B is from a finite element stress
analysis of the poppets, which were made of molded PEEK (polyether
ether ketone) material. At test conditions, the analysis shows that
the impact stress at the poppet head seating surface is 25,700 psi,
which is 177% of the poppet material fatigue strength. This
analysis predicts that the poppets will have a noticeably short
service life when used at such high (greater than 1000 rpm)
operating speeds.
[0035] The impact stresses can be reduced by using smaller, lighter
weight poppets, which has become a common practice, as represented
by the 0.500 in. diameter cylindrical poppets compared with
Sperry's original 1.12 or 1.38 in. diameter mushroom-head poppets.
Nevertheless, it is notable that the impact stress on poppets of
any specific size and mass operating at 1000 rpm is more than 6.25
times the impact stress on the same poppets operating at 400 rpm,
and, when operating at 1800 rpm, the impact stress will be 20.25
times the impact stress at 400 rpm. As a result, size and mass
reduction of poppet proportions to limit the impact stresses to
acceptable levels is generally not practical for high-speed
compressors.
[0036] In light of the above discussion it can be appreciated that
compressor speed (i.e. speeds of 1000 rpm and higher) is a
significant limiting factor in the application of prior art
variable clearance systems. Improvements are therefore necessary
and desirable to make such prior art systems applicable to
high-speed compressors.
[0037] Limitation 2. FIG. 5A through FIG. 5F, which have been
previously described above, illustrate theoretical P-V diagrams for
prior art (e.g. Sperry) variable clearance volume systems. FIG. 7
shows an actual P-V diagram from a compressor operating with such a
prior art system. Notably, there is a delay evident in the
compression stroke, caused when the head ends of the poppets 13
(see FIGS. 3 and 4) do not move off of the valve seat 11 until the
cylinder pressure is higher than the control pressure setting. This
delays the communication time of the clearance pocket volume 15
with the cylinder volume 14 and results in a shorter time of
communication.
[0038] Similarly, there is a delay evident in the expansion stroke,
caused when the stem ends 18 of the poppets 13 do not move off of
the guard seat 12 until a cylinder pressure is reached that is
lower than the control pressure setting. This extends the
communication time of the pocket volume 15 with the cylinder volume
14 and results in a longer time of communication. These delays
change the intended effectiveness of the clearance pocket volume in
unloading the compressor. If the delays are small, the effect can
be limited by setting the control pressure intentionally at a
different pressure, which can somewhat compensate for the defect.
Nevertheless, for reliable and predictable application of the
system, and to avoid damage to the poppets at high operating
speeds, it is necessary to minimize, if not completely eliminate,
such delays.
[0039] It is believed by the present inventors that the cause for
the delay in opening and closing the poppets between the cylinder
volume and the clearance pocket volume is caused by the contact, or
sealing, area between the poppets and their stationary seats when
the poppets are held in contact with the seats. Referring to FIG.
8, the sealing of a poppet 13 to its stationary seat 20, either in
the valve seat or the valve guard, is accomplished, in this case,
by the mating of two conical surfaces, wherein the angle 25 of the
conical surface 23 of the poppet 13 is slightly steeper than the
angle 26 of the conical surface 22 of the stationary seat 20, such
that a narrow circular contact line 24 occurs where surfaces 22 and
23 meet.
[0040] Referring now to FIG. 9, the sealing of a poppet 13 having a
circular surface 32 to its stationary seat 20, either in the valve
seat or the valve guard, can also be accomplished by the mating of
the circular surface 32 with a conical stationary seat surface 22
so that a narrow circular contact line 24 occurs where surfaces 22
and 32 meet. Essentially perfect radial alignment between the two
mating surfaces is required for complete sealing without
leakage.
[0041] Additionally, in practice, due to deformation and surface
wear, the narrow circular contact line 24 has a certain contact
width 40 (and 41) as shown in FIGS. 11A and 11B. This results in a
small annular area on the poppet surface in which no pressure acts
when the poppet is in contact with its stationary seat. During part
of the compression cycle, as shown in FIG. 11A, the control
pressure (P.sub.cont) 46 acts on the stem end 43 of the poppet and
on the head end 42 of the poppet up to the outer edge of the
sealing area (A.sub.2) 44, while cylinder pressure (P.sub.cyl) 47
acts on the head end 42 of the poppet out to the inner edge of the
sealing area (A.sub.1) 45. Also referencing FIG. 3, since area
(A.sub.2) 44 is slightly larger than area (A.sub.1) 45, the
cylinder pressure (P.sub.cyl) 47 must be higher than the control
pressure (P.sub.cont) 46 before the poppet 13 will move off the
valve seat 11 to open communication of the cylinder volume 14 with
the clearance pocket volume 15; i.e., the head end 42 of the poppet
13 moves off the valve seat 11 when
P.sub.cyl>P.sub.cont.times.A.sub.2/A.sub.1.
[0042] During a different part of the compression cycle, as shown
in FIG. 11B, the cylinder pressure (P.sub.cyl) 47 acts on the head
end 42 of the poppet and on the stem end 43 up to the outer edge of
the sealing area (A.sub.4) 48, while the control pressure
(P.sub.cont) 46 acts on the stem end 43 of the poppet out to the
inner edge of the sealing area (A.sub.3) 49. Also referencing FIG.
4, since (A.sub.4) 48 is slightly larger than (A.sub.3) 49, the
cylinder pressure (P.sub.cyl) 47 must be less than the control
pressure (P.sub.cont) 46 before the poppet will move off the guard
seat 12 to stop communication of the cylinder volume 14 with the
clearance pocket volume 15; i.e., the stem end 43 of the poppet
moves off the guard seat 12 when
P.sub.cyl<P.sub.cont.times.A.sub.3/A.sub.4.
[0043] FIG. 12 illustrates this effect for poppets having an inner
circular sealing diameter ranging from 0.360 in, to 0.485 in., and
with an annular sealing width (40 or 41 in FIGS. 11A/B) of 0.003
in. For reference, the inner circular sealing diameter is the
diameter corresponding to areas (A.sub.1) 45 and (A.sub.3) 49 in
FIGS. 11A/B. In FIG. 12, the differential between cylinder pressure
and control pressure is plotted vs. control pressure. For the
compression stroke, the differential is positive (i.e., cylinder
pressure>control pressure) and for the expansion stroke, the
differential is negative (i.e., cylinder pressure<control
pressure). FIG. 12 shows that, for a specific sealing width, a
larger inner circular sealing diameter reduces the differential
pressure.
[0044] FIG. 13 illustrates the influence of the width of the
annular sealing area (40 or 41 in FIGS. 11A/B) for a poppet having
an inner circular sealing diameter of 0.410 in. These calculations
illustrate the importance of a very narrow circular or annular area
of contact and the avoidance of designs or operating conditions
that result in significant wear that would widen the contact area.
FIG. 13 shows that, for a specific inner circular sealing diameter,
a smaller sealing width reduces the differential pressure. In light
of this, it becomes apparent that basic design factors and rules
are desirable in order to limit or minimize the differential
between cylinder pressure and control pressure; such designs could
be used to mitigate and minimize the limitations of prior art (e.g.
Sperry) variable clearance systems.
[0045] Limitation 3. Referring back to FIG. 8, which shows sealing
of a poppet 13 to its stationary seat 20 by the mating of two
conical surfaces, perfect radial alignment is required between the
conical sealing face 23 of the poppet and the conical surface 22 of
the stationary seat, such that the centerline 30 of the poppet and
the centerline 29 of the stationary seat must be perfectly aligned,
resulting in a complete circular contact line 24 between the two
conical surfaces. When there is a radial misalignment of the
centerline 30 of the poppet and the centerline 29 of the stationary
seat, as seen in FIG. 10, contact of the poppet sealing face 23
with the seat surface 22 is imperfect, with contact only made at
one point 33. A gap 31 occurs at other points around the sealing
periphery. Although the poppet axis or centerline 30 may be tilted
to seek better alignment of the sealing face with the seat, the
sealing will nevertheless be imperfect, such that there will be
some leakage of gas past the sealing face while the poppet is in
contact with the seat. This situation occurs on the head end of the
poppet contacting the valve seat, and also on the stem end of the
poppet contacting the guard seat. It also occurs when the poppet
has a circular sealing face of the type shown in FIG. 9.
[0046] The gas control pressure (P.sub.c, P.sub.cont) is higher
than the cylinder pressure (P.sub.cyl) when the head ends of the
poppets 13 are held against the valve seats 11 by the control
pressure (i.e., so that the clearance pocket volume 15 does not
communicate with the cylinder volume 14 (see FIG. 3). Thus, if a
poppet head does not seal perfectly with its valve seat, gas leaks
from the control pressure chamber into the compressor cylinder.
This leakage decreases the control pressure unless it can be
rapidly, i.e. immediately in real time, maintained by the control
valve/pressure regulator 16. Such a decrease in the control
pressure causes the head ends of the poppets to move away from the
valve seat 11 earlier than intended during the compression stroke,
causing the cylinder volume 14 to communicate with the clearance
pocket volume 15 sooner, and for a longer time than intended.
[0047] As FIG. 4 illustrates, the cylinder pressure is higher than
the gas control pressure when the stem ends 18 of the poppets 13
are held against the guard seats 12 by the cylinder pressure (i.e.,
so that the cylinder volume 14 communicates with the clearance
pocket volume 15). Thus, if a poppet stem 18 does not seal
perfectly with the guard seat 12, cylinder pressure can leak into
the control pressure chamber. This leakage increases the control
pressure unless it can be rapidly (i.e., immediately in real time)
maintained by the control valve/pressure regulator 16. Such an
increase in the control pressure during the expansion stroke causes
the stem ends 18 of the poppets to move away from the guard seat 12
sooner than intended, and the opposite, head ends of the poppets 13
to contact the valve seat 11 sooner than intended, terminating the
communication of the clearance pocket volume 15 with the cylinder
volume 14 earlier than intended.
[0048] In both cases of leakage described above, since the control
pressure is different than the intended setting, the control of the
unloader is negatively affected, making its performance less
predictable, more erratic and, therefore, unreliable. In addition,
the early termination of communication of the cylinder volume 14
with the clearance pocket volume 15 during the expansion stroke
means that the compressor is unloaded less than intended. The
amount of leakage is influenced by misalignment of the poppet ends
with their respective seats, the magnitude of the pressure
difference between the cylinder pressure and the control pressure,
and by the speed of the compressor (i.e. higher rpm's or a faster
compressor speed makes it more difficult for the control
valve/regulator 16 to maintain the control pressure in the event of
such leakage).
[0049] Since the time for leakage increases as the control pressure
is decreased (for more unloading), the effect of leakage on the
control pressure increases at lower control pressures. If the
control pressure cannot be instantaneously and continuously
maintained at the intended setting by the control valve/regulator
16, then it cannot be set as low as intended. This reduces the
effectiveness of the unloading, and limits the maximum unloading
potential of the system. Although increasing the volume and line
sizes of the control pressure chamber 17 and the connected system
is one way of partially mitigating the effects of poppet leakage,
such leakage reduces the overall compression efficiency and
increases the operating temperature of the unloader. This
phenomenon is time dependent, meaning that the required response
time of the control system pressure for a high-speed compressor is
much shorter than the required response time for a slow-speed
compressor. Therefore, as compressor speed increases, poppet
leakage requires that the control pressure regulating system must
react increasingly faster in order to maintain the control pressure
appropriately close to the intended setting. This becomes
increasingly impractical as compressor speeds increase, becoming
marginally practical at a compressor speed of about 500 rpm,
unreliable at a compressor speed of about 750 rpm, and essentially
impractical at a compressor speed of about 1000 rpm or higher.
[0050] A further consideration is that high-speed compressors
generally have shorter strokes and, therefore, smaller cylinder
swept volumes than slow-speed compressors producing the same
capacity. Therefore, the effect of communication with a clearance
volume of a specific magnitude is more pronounced for a high-speed
compressor than for a slower-speed compressor having a larger
cylinder swept volume. Thus, deviations between the actual and the
intended control pressure on high-speed compressors causes larger
errors between the actual unloading and the intended unloading that
occurs. This is a further limitation in the application and use of
the prior art Sperry variable clearance system with high-speed
compressors. Therefore, minimization, and ideally elimination, of
poppet leakage is a necessary requirement of any improvement of
these systems for use with high-speed compressors.
[0051] The fundamental cause of leakage is misalignment between the
poppets and the seats. So, when alignment is not perfect, there
will be incomplete sealing, and therefore leakage between the
surfaces of the poppet seats and the mating stationary seats. U.S.
Pat. No. 8,070,461 to Sanford noted this fact, and proposed the use
of poppet sealing rings as a means of reducing the leakage caused
by the misalignment between poppets and seats. However, the sealing
rings are very small, requiring high precision and special tooling
that adds manufacturing complexity and cost. And the requirement
for ring grooves in the outer diameter of the poppets results in
higher stresses in the barrel or wall of the poppets, due to stress
concentration and thinner wall section. Generally, for high-speed
compressors, it is necessary to use a poppet design that is hollow,
closed off on only one end, to minimize the mass of the poppet. The
ring grooves provided by Sanford limit the amount of material that
can be removed from the center of the poppet. These factors can
limit the application range, strength, and the reliability of the
poppets when operating at high speeds. Accordingly, other means are
necessary and desirable for reducing or eliminating leakage between
the poppets and the stationary seats.
[0052] In light of the discussion above, it is apparent that it
would be useful to provide improvements to prior art unloader
systems for application with modern large, high-speed reciprocating
compressors, and specifically for reducing or eliminating poppet
damage and failure, as well as for reducing or eliminating gas
leakage between the poppets and the stationary seats.
SUMMARY OF THE INVENTION
[0053] Accordingly, the present invention provides significant
improvements to known unloader valve assemblies to enable
effective, efficient, and reliable application and use with current
high-speed reciprocating compressors. Specific poppet and seat
design criteria are defined, as well as specific poppets, poppet
seals, cushioning seat plates, and flow area criteria which can
significantly reduce poppet leakage, reduce the control pressure
offset, and reduce the high impact stresses associated with use on
modern high-speed reciprocating compressors.
[0054] A first aspect of the invention provides an unloader valve
assembly for a reciprocating compressor, comprising: (a) a
clearance pocket having a fixed volume; (b) a support dome
including a plurality of support dome ports, wherein each of the
plurality of support dome ports communicates with the clearance
pocket; (c) a valve guard including: (i) a plurality of poppet
guide recesses; (ii) a plurality of poppets, wherein each of the
plurality of recesses houses one of the plurality of poppets, each
of the plurality of poppets having an upper sealing surface and a
lower sealing surface; (iii) a plurality of valve guard seating
surfaces for contacting the upper sealing surfaces of the plurality
of poppets, wherein each of the plurality of valve guard seating
surfaces contacts an upper sealing surface of one of the plurality
of poppets; (iv) a spacer plate portion comprising a plurality of
posts for supporting the valve guard against the support dome,
wherein the spacer plate and the support dome form a control
pressure chamber, wherein each of the plurality of posts include a
central port which communicates with one of the plurality of
support dome ports; (v) a plurality of control pressure ports, each
of the plurality of control pressure ports located in the spacer
plate portion between the plurality of posts, wherein each of the
plurality of control pressure ports communicates with the control
pressure chamber and with one of the plurality of poppet guide
recesses; (d) a valve seat including: (i) a valve seat cushioning
plate for reducing the magnitude of impact stresses acting on the
head end of the poppet, the valve seat cushioning plate including a
plurality of valve seat seating surfaces for contacting the lower
sealing surfaces of the plurality of poppets, wherein each of the
plurality of valve seat seating surfaces contacts a lower sealing
surface of one of the plurality of poppets; and (ii) a plurality of
valve seat ports, wherein each of the plurality of valve seat ports
communicates with a cylinder of a reciprocating compressor.
[0055] A second aspect of the invention provides a control chamber
spacer plate for use with a valve assembly of a reciprocating
compressor unloader, the spacer plate comprising: a plurality of
posts for supporting the spacer plate against a valve support dome
of a valve assembly of a reciprocating compressor unloader to
create a volume between the spacer plate and the valve support
dome, wherein the volume between the spacer plate and the valve
support dome forms a control pressure chamber within the unloader,
wherein each of the plurality of posts include a central port which
communicates with one of a plurality of ports in the valve support
dome, and wherein each of the plurality of ports in the valve
support dome communicates with a clearance pocket volume of the
unloader; and a plurality of control pressure ports, each of the
plurality of control pressure ports located in the spacer plate
between the plurality of posts, wherein each of the plurality of
control pressure ports communicates with the control pressure
chamber.
[0056] A third aspect of the invention provides a pressure breaker
poppet for use with an unloader valve assembly of a reciprocating
compressor, the pressure breaker poppet comprising: a head end
having a first diameter, the head end including: an upper sealing
surface for contacting a seating surface of a valve guard of an
unloader valve assembly of a reciprocating compressor, wherein the
valve guard includes a control pressure port connected to a
clearance pocket of the unloader valve assembly; and a lower
sealing surface for contacting a seating surface of a valve seat of
the unloader valve assembly, wherein the valve seat includes a
valve seat port connected to a cylinder volume of the compressor; a
stem end having a second diameter which is smaller than the first
diameter of the head end, the stem end including a plurality of
steps radially protruding from the second diameter; and a sealing
guide stem having a third diameter which is smaller than the second
diameter of the stem end, wherein the sealing guide stem is sized
to fit within the control pressure port in the valve guard.
[0057] A fourth aspect of the invention provides a pressure breaker
poppet for use with an unloader valve assembly of a reciprocating
compressor, the pressure breaker poppet comprising: a head end
having a first diameter, the head end including: an upper sealing
surface for contacting a seating surface of a valve guard of an
unloader valve assembly of a reciprocating compressor, wherein the
valve guard includes a control pressure port connected to a
clearance pocket of the unloader valve assembly; and a lower
sealing surface for contacting a seating surface of a valve seat of
the unloader valve assembly, wherein the valve seat includes a
valve seat port connected to a cylinder volume of the compressor;
and a stem end having a second diameter which is smaller than the
first diameter of the head end, the stem end including a plurality
of steps radially protruding from the second diameter.
[0058] A fifth aspect of the invention provides an impact tolerant
self-sealing poppet for use with an unloader valve assembly of a
reciprocating compressor, the impact tolerant self-sealing poppet
comprising: a head end having an outer diameter, the head end
including a plurality of steps radially protruding from the outer
diameter of the head end and a lower sealing surface for contacting
a seating surface of a valve seat of an unloader valve assembly of
a reciprocating compressor; and a stem end having an outer diameter
that is substantially the same as the outer diameter of the head
end, the stem end including a plurality of steps radially
protruding from the outer diameter of the stem end and an upper
sealing surface for contacting a seating surface of a valve guard
of the unloader valve assembly.
[0059] A sixth aspect of the invention provides an impact tolerant
self-sealing floating seat poppet for use with an unloader valve
assembly of a reciprocating compressor, the impact tolerant
self-sealing floating seat poppet comprising: a head end having an
outer diameter, the head end comprising a head end piece including
a central hollow portion having an inner diameter, and a lower
sealing surface for contacting a seating surface of a valve seat of
an unloader valve assembly of a reciprocating compressor; a stem
end having an outer diameter that is substantially the same as the
outer diameter of the head end, the stem end comprising a stem end
piece including a central hollow portion having an inner diameter
that is substantially the same as the inner diameter of the head
end piece, and an upper sealing surface for contacting a seating
surface of a valve guard of the unloader valve assembly; a central
core positioned between the head end piece and the stem end piece,
the central core comprising a hollow portion, a plurality of steps
radially protruding from the outer diameter of the central core,
and bulbous-shaped ends for connecting to the head end piece and
the stem end piece, wherein the bulbous-shaped ends have a larger
outer diameter than the inner diameters of the head end piece and
the stem end piece.
[0060] A seventh aspect of the invention provides a diaphragm seal
poppet for use with an unloader valve assembly of a reciprocating
compressor, the diaphragm seal poppet comprising: a head end having
an outer diameter and including a lower sealing surface for
contacting a seating surface of a valve seat of an unloader valve
assembly of a reciprocating compressor; a stem end having an outer
diameter that is substantially the same as the outer diameter of
the head end; and a flexible diaphragm seal projecting from the
stem end, the diaphragm seal comprising: a bulbous outer diameter
clamped into a recess of a valve guard of the unloader valve
assembly; a seating recess for contacting a seating surface of the
valve guard; and at least one strain relief loop positioned between
the bulbous outer diameter and the seating recess.
[0061] The nature and advantages of the present invention will be
more fully appreciated from the following drawings, detailed
description, and claims.
BRIEF DESCRIPTION OF THE DRAWINGS
[0062] The accompanying drawings illustrate embodiments of the
invention and, together with a general description of the invention
given above, and the detailed description given below, serve to
explain the principals of the invention.
[0063] FIGS. 1A-1F illustrate the steps of the processes inside a
reciprocating compressor cylinder without (FIGS. 1A-1D) and with
(FIGS. 1E-1F) a fixed clearance volume communicating with the
volume inside the compressor cylinder;
[0064] FIG. 2 illustrates a theoretical pressure-volume diagram for
a reciprocating compressor cylinder with and without a fixed
clearance volume communicating with the volume inside the
compressor cylinder;
[0065] FIG. 3 illustrates a cross-section of a prior art (i.e.
Sperry) variable clearance system with the system inactive, such
that the fixed clearance volume does not communicate with the
volume inside the compressor cylinder;
[0066] FIG. 4 illustrates a cross-section of the Sperry variable
clearance system with the system active, such that the fixed
clearance volume communicates with the volume inside the compressor
cylinder;
[0067] FIGS. 5A-5F illustrate a representation of theoretical
pressure-volume diagrams showing the effect of the Sperry variable
clearance system to provide several increments of unloading,
ranging from no unloading to maximum unloading for a specific
example;
[0068] FIG. 6A illustrates failed poppets of a prior art variable
clearance system for high-speed reciprocating compressors;
[0069] FIG. 6B illustrates calculated poppet impact stresses which
typically result from application of prior art variable clearance
system for high-speed reciprocating compressors;
[0070] FIG. 7 illustrates a pressure-volume diagram for a prior art
variable clearance system for high-speed reciprocating
compressors;
[0071] FIG. 8 illustrates a conical poppet seating surface engaging
a conical stationary valve seat;
[0072] FIG. 9 illustrates a circular poppet seating surface
engaging a conical stationary valve seat;
[0073] FIG. 10 illustrates a conical poppet seating surface
engaging a conical stationary valve seat with misalignment of the
poppet and stationary seat centerlines;
[0074] FIG. 11A illustrates the pressure distribution acting on a
poppet when held against the valve seat;
[0075] FIG. 11B illustrates the pressure distribution acting on a
poppet when held against the guard seat;
[0076] FIG. 12 illustrates a graphical representation of the
differential between cylinder pressure and control pressure as a
function of poppet sealing diameter for a fixed seat sealing
width;
[0077] FIG. 13 illustrates a graphical representation of the
differential between cylinder pressure and control pressure as a
function of poppet seat sealing width for a fixed sealing
diameter;
[0078] FIG. 14 illustrates a conical poppet seating surface
separated from a conical stationary valve seat by the lift, or
travel distance, of the poppet;
[0079] FIG. 15A illustrates a cross-section of an unloader assembly
according to the present invention;
[0080] FIG. 15B is a perspective view of the valve seat area of the
unloader assembly encircled in FIG. 15A;
[0081] FIG. 16A illustrates a cross-section of an unloader assembly
according to the present invention with a pressure breaker poppet
having an extended sealing guide stem;
[0082] FIG. 16B is a perspective view of the pressure breaker
poppet of FIG. 16A;
[0083] FIG. 17 is a perspective view of pressure breaker poppet
without an extended sealing guide stem;
[0084] FIG. 18A illustrates a cross-section of a valve seat area of
an unloader assembly including an impact tolerant self-sealing
poppet;
[0085] FIG. 18B is a perspective view of the self-sealing poppet of
FIG. 18A;
[0086] FIG. 19 is a perspective view of a self-sealing floating
seat poppet;
[0087] FIG. 20 is a perspective view of a diaphragm seal
poppet;
[0088] FIG. 21 illustrates a cross-section of the present invention
showing a passive valve seat cushioning plate with a cushioning pad
element;
[0089] FIG. 22A illustrates a passive valve seat cushioning plate
with an o-ring cushioning element, showing the poppets held against
the valve guard;
[0090] FIG. 22B illustrates a passive valve seat cushioning plate
with an o-ring cushioning element, showing the poppets held against
the valve seat plate;
[0091] FIG. 23A illustrates a passive valve seat cushioning plate
with a spring cushioning element, showing the poppets held against
the valve guard;
[0092] FIG. 23B illustrates a passive valve seat cushioning plate
with a spring cushioning element, showing the poppets held against
the valve seat plate;
[0093] FIGS. 24A-24D illustrate test data demonstrating the
improvements of the present invention over similar prior art
systems.
DETAILED DESCRIPTION OF THE INVENTION
[0094] The present invention provides improvements to prior art
unloader valve assemblies for use with reciprocating compressors to
allow selective communication between the compressor cylinder and a
clearance pocket, as detailed above. The invention can provide
improved communication between the compressor cylinder and the
clearance pocket, and discloses specific poppet and valve seat
design criteria, flow area criteria, poppet seals, and cushioning
seat plates that can reduce or eliminate poppet leakage, the
control pressure offset, and the high impact stresses on the
poppets that are associated with operating prior art unloader
systems at high operating speeds. The embodiments disclosed herein
provide an effective, efficient, and reliable solution for
application of prior art unloader systems at compressor speeds of
1000 rpm or higher.
[0095] Poppet Seat Design Requirements--Novel poppets (and mating
valve seats) are disclosed herein, which are useful for maximizing
unloader efficiency and achieving acceptable poppet service life.
The poppet designs (also referred to in the prior art as a poppet
valve member, a closing element, or a valve poppet) disclosed
herein are a result of the optimization of several desirable
geometric features, for example, the poppet's mass (and therefore
its physical size) must be sufficiently small so as to limit the
impact stresses on the poppet and valve seat faces at compressor
speeds higher than 1000 rpm and up to 1800 rpm. The diameter of the
port that is sealed by the poppet must be sufficiently large so as
to minimize pressure losses as gas flows rapidly between the
compressor cylinder volume and the clearance pocket volume during
each stroke of the compressor piston. The diameter of the narrow
circular contact line between the poppet sealing face and its
stationary valve seat must be sufficiently large as practical, and
the annular width of the circular contact line must be sufficiently
narrow as practical to minimize the differential pressure required
between cylinder pressure and control pressure for opening and
closing the poppets. The diameter and the width of the circular
contact line must be sufficiently large so as to limit the impact
stress on the poppet seat face when the poppet closes rapidly
against the mating face in the stationary valve seat. The inner
diameter of the circular contact line must be sufficiently larger
than the port diameter to ensure that the circular contact line
does not overlap the port and cause leakage. The outer diameter of
the circular contact line must be sufficiently smaller than the
poppet outside diameter to limit the corner stress caused by the
poppet closing against the stationary seat. The poppet lift (travel
distance between the seat and the guard within the valve assembly)
must be sufficiently large so as to minimize pressure losses as gas
flows rapidly back and forth between the compressor cylinder volume
and the clearance pocket volume, but the lift must also be
sufficiently small so as to the limit the impact stresses on the
valve seat faces at speeds higher than 1000 rpm and up to 1800
rpm.
[0096] In light of the geometric features described above which are
desirable for maximizing unloader efficiency and achieving
acceptable poppet service life, the present invention discloses
several poppet and valve seat parameters useful for optimizing
unloader efficiency and reliability: (1) The poppet preferably has
an outside diameter of no larger than 0.510 in. and no smaller than
0.490 in., more preferably no larger than 0.502 in. and no smaller
than 0.498 in., and most preferably no larger than 0.500 in. and no
smaller than 0.499 in.; (2) The conical angle of a stationary seat
(referenced as 26 in FIG. 8) is no larger than 45 degrees and no
smaller than 35, preferably no larger than 42.5 degrees and no
smaller than 37.5, and most preferably 40 degrees; (3) For poppets
having a conical poppet sealing face, the angle of the poppet seat
(referenced as 25 in FIG. 8) is preferably no less than 1.0 degree
and no more than 5.0 degrees larger than the conical angle of the
mating stationary seat (referenced as 26 in FIG. 8), and more
preferably 3.0 degrees larger than the conical angle of the
stationary seat; (4) For poppets having a circular seat (as of the
type shown in FIG. 9), the circular seat is designed to maintain
the preferred sealing diameter requirement (see item 6, below) and
to maintain acceptable impact stress in the poppet seat. The
circular seat may be formed by a single radius, a blend of multiple
radii, or a combination of one or more conical angles and one or
more radii; (5) The port diameter (referenced as 27 in FIG. 8 and
FIG. 9), is preferably no smaller than 0.380 in. and no larger than
0.410 in., and more preferably 0.400 in.; (6) The center or mean
diameter of the narrow circular contact line between the conical
poppet seat (referenced as 24 in FIG. 8, or the circular poppet
seat 24 in FIG. 9) and the conical stationary seat that is
preferably no less than 0.020 in. and no more than 0.045 in. larger
than the port diameter, more preferably no less than 0.030 in. and
no more than 0.044 in. larger than the port diameter, and most
preferably no less than 0.038 in. and no more than 0.042 in. larger
than the port diameter; (7) The poppet traveling distance, also
referred to and labeled as the lift 52 in FIG. 14, is such that the
lift area (which is calculated by multiplying the lift by the
circumference of the inner diameter of the circular contact line
between the poppet sealing face and the stationary conical seat) is
no less than 97% and no more than 103% of the port area 50 in FIG.
14; and (8) For poppets having a conical poppet seat, the
intersection of the conical surface and the outer diameter of the
poppet is rounded with a radius of no less than 0.015 in., to
reduce pressure drop caused by the corner and to avoid an otherwise
sharp, weak corner that would chip or break off with repeated
impact.
[0097] Flow Path Design Requirements--The rapid, cyclical flow
occurring back and forth between the compressor cylinder volume and
the clearance pocket volume with each stroke of the compressor
piston should be as unrestricted as practical to maximize
effectiveness and minimize pressure losses and resulting parasitic
power losses. Minimization of flow restrictions is also desirable
in order to minimize the differential between the cylinder pressure
and the control pressure, once flow to and from the compressor
cylinder volume occurs, to prevent fluttering (i.e., repeated
unintended opening and closing) of the poppets. Fluttering of the
poppets can increase parasitic power losses and reduce the
reliability of the poppets and, therefore, of the unloader.
[0098] FIG. 15A shows a cross-section of one embodiment of an
unloader assembly for a reciprocating compressor, according to the
present invention. The unloader assembly includes an outer head 212
and an inner head 67, which are connected by retention bolts
passing through holes 224. Together the outer head 212 and the
inner head 67 form a clearance pocket volume 15 which is fluidly
connected to the cylinder volume 14 of a reciprocating compressor.
An outer control pressure port 217 passes through the outer head
212 (via a retainer post 218) and an unloader valve support dome 69
to reach an inner control pressure chamber 17 at one end. The other
end of the outer control pressure port 217 is connected to an
external pressure control valve or pressure regulator (not shown),
as is known in the art. A retainer cap 220 is secured to the other
head 212 via retainer cap bolts 222 so as to compress the retainer
post 218 such that it clamps the support dome 66 and the rest of
the internal assembly tightly against the sealing gasket 300 on the
shoulder 301 within the inner head 67.
[0099] Referring to FIGS. 15A and 15B, the path for the cyclic flow
includes the entrances 60 from the compressor cylinder volume 14 to
the multiple ports 19 in the valve seat 11, the aggregate area of
the multiple ports 19 in the valve seat, the aggregate lift area 61
of the multiple poppets, the aggregate passage area 62 between the
valve seat cushioning plate 68 and the valve guard 65 (or between
the valve seat 11 and the valve guard 65 if no cushioning plate 68
is utilized), and the sum of the aggregate area of the multiple
ports 66 in the unloader valve support dome 69 and the passage area
63 between the outer diameter of the valve guard 65 and the inner
diameter of the head 67 to the clearance pocket volume 15; and all
the way back again from the clearance pocket volume 15 to the
compressor cylinder volume 14. The lift area 61 of the poppets is
normally the limiting flow area, as it relates to the number of
poppets and to the magnitude of the lift or opening distance of the
poppets during operation. The number of poppets depends on the
quantity that can be geometrically arranged in the valve seat 11
and the valve guard 65 while maintaining sufficient strength in the
seat and guard. The optimal lift, or travel/opening distance of the
poppet, is also determined by the impact stress on the poppet
during operation at high speed. The maximum clearance pocket volume
is determined by pressure losses resulting from the aforementioned
flow path between the compressor cylinder volume 14 and the
clearance pocket volume 15, including the aggregate poppet lift
area 61.
[0100] The following two design relationships between the
respective flow areas have been determined by the inventors to
optimize efficiency and reliability. First, referring to FIG. 14,
the lift area 51 (defined as the circumference of the inner edge of
the circular contact line 28 between the conical poppet seat 23 or
the circular poppet seat 32 and the conical stationary seat 22
multiplied by the lift or opening/traveling distance 52 of the
poppet 13) is preferably no less than 95% and no more than 110% of
the port area 50 in the stationary seat 20, and more preferably no
less than 100% and no more than 105% of the port area 50 in the
stationary seat 20. Second, the aggregate area of all passages or
restrictive points communicating flow to and from the clearance
pocket volume 15 and the poppet lift areas 51 when the poppets are
open is no less than 125% of the sum of all of the port areas 50 in
the valve seat, and preferably no less than 150% of the sum of all
of the port areas 50 in the valve seat.
[0101] In order to meet the second design requirement noted above,
the specific design shown FIGS. 15A and 15B is presented. This
design may apply to the spacer plate 64 (FIG. 15A) or to a valve
guard 65 which incorporates the features of the spacer plate 64
(FIG. 15B). The specific design feature is the incorporation of
multiple posts 72, each post having a truncated triangular shape.
The multiple posts 72 separate and support the spacer plate 64 or
the valve guard 65 (shown in FIG. 15B incorporating the features of
the spacer plate 64) against the unloader valve support dome 69,
with the volume between the spacer plate 64 and the unloader valve
support dome 69 forming the control pressure chamber 17. Looking at
FIG. 15B, each post 72 has a central port 70 that aligns and
communicates with a port 66 (FIG. 15A) in the unloader valve
support dome 69 that communicates with the clearance pocket volume
15, such that the aggregate area of the ports 70 is additive to the
passage area 63 between the outer diameter of the valve guard 65
and the inner diameter of the cylinder head 67 to the clearance
pocket volume 15. Separate control pressure ports 71 are located
between the multiple posts 72, and these ports 71 are aligned with
their individual poppet guide recesses 85 to provide communication
of gas from the control pressure chamber 17 to the stem ends of
their respective poppets. The truncated triangle shape of each of
the multiple posts 72 optimizes the support area provided by the
posts and the flow area around the posts, while also enabling
efficient machining by allowing milling of the posts 72 and the
control pressure chamber 17 from a solid steel plate. Other shapes
may be used depending on the manufacturing processes, the number of
ports, and other specific requirements of applications.
[0102] Improved Poppet Designs--For successful application of the
variable clearance system in high-speed compressors, improved
poppet designs are provided by the present invention that can
address the aforementioned limitations associated with high impact
forces, stresses, and leakage past the poppet stems into and out of
the control pressure chamber. Depending on the specific application
and operating conditions of the reciprocating compressor system,
several alternative designs incorporating some or all of the
described features can be incorporated into the poppet design.
[0103] A) Pressure Breaker Poppet with Extended Sealing Guide
Stem--Referring to FIGS. 16A and 16B, a pressure breaker poppet 74
includes a head end 75 having a first diameter, a stem end 76
having a second diameter that is smaller than the first diameter,
and a sealing guide stem 77 having a third diameter which is
smaller than the second diameter. This design results in the poppet
74 having a smaller mass than a prior art solid poppet with an
overall diameter the same as the head end. The poppet 74 may be
solid or hollow, containing an internal volume 78 which is based on
an optimization of the requirements for high strength and low mass,
in order to limit the impact stresses on the head end 75. The
poppet material may be carbon-filled PEEK (polyether ether ketone),
glass-filled PEEK or another high-strength, non-metallic
material.
[0104] In the "closed" position, as shown in FIG. 16B, a lower
sealing surface 79 of the poppet head end 75 is held in contact
with the conical surface 80 of the valve seat cushioning plate 68
(or valve seat 11 if no cushioning plate is used). This occurs when
the control pressure 46 is higher than the compressor cylinder
pressure 47. In contrast, in the "open" position the cylinder
pressure 47 is higher than the control pressure 46, and an upper
sealing surface 81 of the poppet head end 75 is held in contact
with the conical upper seating surface 87 of the valve guard 65. A
control pressure volume 82 remains in the poppet guide recess 85
above the poppet stem 76 when in the closed position, and this
volume 82 provides a cushioning effect as the poppet opens and
moves towards the seating surface 87 of the valve guard 65, without
creating excessive delay or parasitic energy loss.
[0105] Two or more pressure breaker steps 83 can be provided as
radial protrusions projecting from the outer diameter of the stem
end 76 of the poppet. These steps 83 interrupt the path of any
leakage flow around the stem end, i.e. through the annular space 84
between the poppet stem 76 and the poppet guide recess 85 to and
from the control pressure chamber 17 and the compressor cylinder 14
or the clearance pocket volume 15. The pressure breaker steps 83
may be radial protrusions with square edges as shown, or they may
be radial protrusions with sharp edged labyrinth teeth, slanted
protrusions with labyrinth teeth, or other geometric profiles or
any combination of profiles. The outer diameter of the pressure
breaker steps 83 is sized to fit inside the bore of the poppet
guide recess 85 with minimal radial clearance. However, sufficient
radial clearance is still provided so as not to interfere with the
alignment of the lower sealing surface 79 of the poppet head end 75
with the conical surface 80 of the valve seat cushioning plate 68
(or valve seat 11, if no cushioning plate is used) upon closure, or
with alignment of the sealing surface 81 of the poppet stem end 76
with the conical seating surface 87 of the valve guard 65 upon
opening.
[0106] The change in diameter between the poppet guide recess 85
and the control pressure port 86 interrupts the path of any leakage
to and from the control pressure chamber 17 and the compressor
cylinder 14 or the clearance pocket volume 15. This interruption
can create localized eddies and turbulence, which can create a
localized pressure drop that interrupts, reduces and minimizes the
rate of leakage. The annular space 88 between the outer diameter of
the sealing guide stem 77 and the diameter of the control pressure
port 86 is smaller and therefore more restrictive than the annular
space 84 between the stem end 76 of the poppet and the poppet guide
recess 85, and provides a further interruption in the path of any
leakage to and from the control pressure chamber 17 and the
compressor cylinder 14 or the clearance pocket volume 15. When the
poppet 74 opens and closes, the aerodynamic frictional drag created
by the pressure breaker steps 83 limits the velocity of the poppet
as it moves across the gap from the valve guard 65 to the
cushioning plate 68 (or to the valve seat 11 if no cushioning plate
is used). This beneficial effect reduces the impact velocity and
therefore the impact stress on the lower sealing surface 79 of the
poppet head.
[0107] B) Pressure Breaker Poppet Without Extended Sealing Guide
Stem--Referring to FIG. 17, a simplified version of the pressure
breaker poppet 74 of FIGS. 16A and 16B is illustrated. This poppet
174 does not include an extended sealing guide stem ([77, see FIG.
16), and, although this version provides less resistance and/or
interruption of leakage to and from the control pressure chamber
17, it has less mass and higher strength than the version of FIG.
16, which is useful for applications requiring both high speed and
high compression ratios. The poppet 174 may be solid or hollow,
containing an internal volume 78 which is based on an optimization
of the requirements for high strength and low mass in order to
limit the impact stresses on the head end 75. The poppet material
may be carbon-filled PEEK, glass-filled PEEK or another
high-strength, non-metallic material.
[0108] C) Impact Tolerant Self-Sealing Poppet--Referring to FIGS.
18A and 18B, an impact tolerant self-sealing poppet 89 is
illustrated which includes a head end 75 and a stem end 76 in which
the diameter of the head end is substantially the same as the
diameter of the stem end. The poppet 89 may be solid or hollow and
contain an internal volume 78 based on an optimization of the
requirements for high strength and low mass in order to limit the
impact stresses on the head end 75.
[0109] In the "closed" position, as shown in FIG. 18B, the lower
sealing surface 79 of the poppet head end 75 is held in contact
with the conical surface 80 of the valve seat cushioning plate 68
(or directly with the valve seat 11 if no cushioning plate is
used). This occurs when the control pressure 46 is higher than the
compressor cylinder pressure 47. In contrast, in the "open"
position the cylinder pressure 47 is higher than the control
pressure 46, and the sealing surface 81 of the poppet stem end 76
is held in contact with the conical seating surface 87 of the valve
guard 65. A control pressure volume 82 remains in the poppet guide
recess 85 above the poppet stem 76 when in the closed position, and
this control pressure volume 82 provides a cushioning effect as the
poppet opens and moves towards the seating surface 87 of the valve
guard 65, without creating excessive delay or parasitic energy
loss.
[0110] Two or more pressure breaker steps 83 are provided on the
outer diameter of the stem end 76 and two or more pressure breaker
steps 83 are also provided on the outer diameter of the head end 75
of the poppet 89, such that the steps 83 interrupt the path of any
leakage flow through the annular space 84 between the poppet 89 and
the poppet guide recess 85 in the guard 65 to and from the control
pressure chamber and the compressor cylinder or the clearance
pocket volume. The pressure breaker steps may be radial protrusions
with square edges as shown, radial protrusions with sharp edged
labyrinth teeth, slanted protrusions with labyrinth teeth, pressure
activated sealing strips, or they may be other geometric profiles
or any combination of profiles. The outer diameter of each pressure
breaker step 83 is sized to fit inside the bore of the poppet guide
recess 85 with minimal radial clearance. However, sufficient radial
clearance is still provided so as not to interfere with the
alignment of the lower sealing surface 79 of the poppet head end 75
with the conical surface 80 of the valve seat cushioning plate 68
(or valve seat 11, if no cushioning plate is used) upon closure, or
with alignment of the sealing surface 81 of the poppet stem end 76
with the conical seating surface 87 of the valve guard 65 upon
opening.
[0111] The change in diameter between the poppet guide recess 85
and the control pressure port 71 interrupts the path of any leakage
to and from the control pressure chamber 17 and the compressor
cylinder 14 or the clearance pocket volume 15. This interruption
can create localized eddies and turbulence, which can create a
localized pressure drop that interrupts, reduces and minimizes the
rate of leakage. When the poppet 89 opens and closes, the
aerodynamic frictional drag created by the pressure breaker steps
83 limits the velocity of the poppet as it moves across the gap
from the valve seat 68 to the valve guard 65. This beneficial
effect reduces the impact velocity and therefore the impact stress
on the poppet sealing surfaces 79, 81.
[0112] The impact tolerant self-sealing poppet 89 can be made of a
resilient material such as carbon-filled PTFE, which is more impact
resistant than PEEK or other materials. Depending on the
application requirements, this poppet design can also be made of
carbon-filled PEEK.
[0113] D) Impact Tolerant Self-Sealing Floating Seat
Poppet--Referring to FIG. 19 a multi-piece poppet is shown having a
central core 90, the central core including bulbous-shaped ends for
connection to a stem end piece 91 and a head end piece 92. The
central core 90 can include a hollow portion 78 to reduce its mass,
and the bulbous-shaped ends have a larger outer diameter 95 than
the inner diameters 96, 196 of the end pieces 91, 92. During
assembly, special tooling can be employed to elastically compress
the bulbous ends of the central core 90 as they are connected to
the end pieces 91, 92, and the bulbous ends can then recover their
original shape, so that they remain engaged with and retained
inside the two end pieces 91, 92. A radial clearance 93 is
maintained between the inside diameter 96, 196 of each end piece
and the diameter 94 of each stem of the central core 90 and also
between the larger outer diameter 95 of the bulbous ends and the
respective inside diameters 97, 197 of the end pieces. The radial
clearance permits the end pieces 91, 92 to "float" or move radially
to align with the conical surface 80 of the valve seat cushioning
plate 68 (or the valve seat 11 if no cushioning plate is used) and
the conical seating surface 87 of the valve guard 65, as the
central core 90 remains aligned with the poppet guide recess
85.
[0114] In the closed position illustrated in FIG. 19, the lower
sealing surface 79 of the poppet head end 92 is held in contact
with the conical surface 80 of the stationary valve seat cushioning
plate 68 (or the valve seat 11, if no cushioning plate is used).
This occurs when the control pressure 46 is higher than the
compressor cylinder pressure 47. In contrast, in the "open"
position the cylinder pressure 47 is higher than the control
pressure 46, and the sealing surface 81 of the poppet stem end 91
is held in contact with the conical seating surface 87 of the valve
guard 65. A control pressure volume 82 remains in the poppet guide
recess 85 above the poppet stem 76 when in the closed position, and
this volume 82 provides a cushioning effect as the poppet opens and
moves towards the seating surface 87 of the valve guard 65, without
creating excessive delay or parasitic energy loss.
[0115] A plurality of pressure breaker steps 83 (e.g. four, as
shown in FIG. 19) can be provided on the outer diameter of the
central core 90, such that the steps interrupt the path of any
leakage flow through the annular space 84 between the central core
90 of the poppet and the poppet guide recess 85 in the guard 65 to
and from the control pressure chamber 17 and the compressor
cylinder 14 or the clearance pocket volume 15. The pressure breaker
steps 83 may be radial protrusions with square edges as shown,
radial protrusions with sharp edged labyrinth teeth, slanted
protrusions with labyrinth teeth, pressure activated sealing
strips, or they may be other geometric profiles or any combination
of profiles.
[0116] The outer diameter of the pressure breaker steps 83 is sized
to fit inside the bore of the poppet guide recess 85 with minimal
radial clearance. However, there is still sufficient radial
clearance so as not to interfere with the alignment of the lower
sealing surface 79 of the poppet head end 75 with the conical
surface 80 of the valve seat cushioning plate 68 upon closure, or
with alignment of the sealing surface 81 of the poppet stem end 76
with the conical seating surface 87 of the valve guard 65 upon
opening.
[0117] The change in diameter between the poppet guide recess 85
and the control pressure port 86 interrupts the path of any leakage
to and from the control pressure chamber 17 and the compressor
cylinder 14 or the clearance pocket volume 15. This interruption
can create localized eddies and turbulence, which can create a
localized pressure drop that interrupts, reduces and minimizes the
rate of leakage. When the poppet 89 opens and closes, the
aerodynamic frictional drag created by the pressure breaker steps
83 limits the velocity of the poppet as it moves across the gap
from the valve seat 68 to the valve guard 65. This beneficial
effect reduces the impact velocity and therefore the impact stress
on the poppet sealing surfaces 79, 81.
[0118] The central core 90 can be made of a resilient material,
such as carbon-filled PTFE, or a stronger material such as
carbon-filled PEEK, or any other non-metallic material of
sufficient strength and temperature rating. When the central core
is made of a resilient material, it increases the impact resistance
of the poppet. The end pieces 91, 92 can be made of a stronger
non-metallic material such as carbon-filled PEEK, glass-filled
PEEK, or other material of sufficient temperature rating and impact
strength.
[0119] E) Diaphragm Seal Poppet--This poppet includes a head end
and a stem end in which the diameter of the head end is
substantially the same as the diameter of the stem end, and a
diaphragm seal 98, as described below and illustrated in FIG. 20.
The diaphragm seal valve poppet is an alternative embodiment of the
improved poppet designs described above. This design eliminates the
problem of leakage past the poppets, and the resulting effect such
leakage can have on the control pressure. The diaphragm seal poppet
is particularly applicable when the unloader valve assembly is used
in compressors with low molecular weight gases such as hydrogen,
helium or mixtures of gases that contain mostly hydrogen.
[0120] Referring to FIG. 20, a flexible diaphragm seal 98 is
employed over the stem end of each poppet 99 to prevent leakage of
gas between the control pressure 46 in the control pressure port 71
and compressor cylinder pressure 47 in the valve seat port 19 or
the clearance pocket volume. The poppet diaphragm seal 98 has a
bulbous outer diameter 106, similar to an O-ring, that is clamped
by a control chamber spacer plate 64 into a recess 105 of the valve
guard 65. This is useful in preventing the leakage of gas around
the poppet and the seal.
[0121] Since there is no longer a leakage path between the outer
diameter of the poppet and the poppet guide recess 85 in the valve
guard 65, this configuration permits the radial clearance between
the poppet 99 and the poppet guide recess 85 to be large enough to
ensure that the poppet can float radially and align with the
conical surface 80 of the valve seat cushioning plate 68 (or valve
seat 11 if no cushioning plate is used) without contacting the
walls of the poppet guide recess 85 in the valve guard 65. When the
control pressure 46 exceeds the cylinder pressure 47 by an amount
that is proportional to the ratio of the poppet head end seating
area to the area of the poppet stem end, the diaphragm seal 98 is
pushed against the top edge of the poppet 99, holding it against
the conical surface 80 of the valve seat cushioning plate 68, as
shown in FIG. 20 (or valve seat 11 if no cushioning plate is used).
When the cylinder pressure 47 exceeds the control pressure 46 by an
amount that is proportional to the ratio of the area of the poppet
stem end to the poppet head end seating area, the poppet 99 is
pushed against the bottom of the diaphragm seal 98. This pushes the
diaphragm seal 98 across the lift or opening distance 100 until it
rests against the diaphragm seal seating recess 101. The diaphragm
seal 98 incorporates one or more strain relief loops 102 that
reduce the bending stress in the seal as it deflects across the
opening distance 100. The diaphragm seal seating recess 101 is
designed to accommodate a specific amount of opening distance 100
and the elastic deformation of the seal 98 without excessively high
bending stresses occurring in the diaphragm seal 98. The shape of
the diaphragm seal 98 may be as shown in FIG. 20, or it may be any
other shape that serves the same purposes of reliably accommodating
the poppet lift 100 and sealing between the control pressure volume
82 and the poppet guide volume 104.
[0122] Use of the diaphragm seal 98 arrangement as shown in FIG. 20
allows the poppet 99 to have generous radial clearance with the
poppet guide recess 85, making the poppet seating more tolerant of
misalignment of the poppet guide recess 85, the poppet 99, and the
conical surface 80 of the valve seat cushioning plate 68 (or valve
seat 11 if no cushioning plate is used). It also permits somewhat
less precision in the alignment required between the valve seat
cushioning plate 68 (or valve seat 11 if no cushioning plate is
used) and the valve guard 65. The diaphragm seal 98 can be made of
an elastomeric material such as fluorocarbon, ethylene propylene,
nitrile, polyester or polyether urethane,
tetrafluoroethylene-propylene or other similar compounds. The seal
material is preferably selected based on the operating conditions,
amount of elastic stretch required and compatibility with the gas
composition and any contaminants, including lubricants, in the gas
stream.
[0123] Passive Valve Seat Cushioning Plate--Referring back to FIG.
6, it is notable that one problematic characteristic of prior art
poppets is that when a poppet fails, the failure originates in the
lower sealing surface of the poppet head, but not in the stem end
sealing surface. Compressor speeds at or above 1000 rpm can cause
such failed poppets, and this is a significant limiting factor in
the application of prior art variable clearance systems and poppets
to modern high-speed compressors. Improvements are therefore
necessary and desirable to make such prior art systems applicable
to high-speed compressors. For example, the inventive poppet 89 as
illustrated FIGS. 18A and 18B can be positioned in a guide recess
85 that includes a relatively small control pressure volume 82
communicating with the control pressure chamber 17 via the control
pressure port 71 in the valve guard 65. Therefore, as the poppet
moves toward the stem end of the guide recess 85, the gas acts like
a spring as the poppet forces gas through the port 71 and into the
control pressure chamber 17. The compression work done by the
poppet slows the rate of deceleration and provides an effective
cushioning of the poppet as it approaches the guard seat 87.
However, when the poppet changes direction and moves toward the
valve seat cushioning plate 68 (or the valve seat 11, if no
cushioning plate is used), there is comparatively little resistance
between the lower sealing surface 79 of the poppet head 75 and the
conical surface 80 of the stationary cushioning plate 68. This lack
of resistance at the head end of the poppet exists because the gas
can flow through the multiple ports 19 in the valve seat 11 into
the much larger cylinder volume 14, and also into the clearance
pocket volume 15. As a result, the poppet is not decelerated, and
it impacts the seat surface 80 with much higher impact velocity and
force.
[0124] The present invention provides a means for cushioning the
impact of the head ends of the poppets with the stationary valve
seats. Referring to FIG. 21, a valve seat cushioning plate 68 is
oriented over the valve seat 11 and aligned by two or more
precision guide sleeves 110, each of which is retained by a cap
screw 109 and sealed by an O-ring 111, such that ports 108 in the
valve seat cushioning plate 68 align with the ports 19 in the valve
seat 11. The valve seat cushioning plate 68 can be constructed of a
non-metallic material, such as carbon-filled PEEK or glass-filled
PEEK, which is more resilient than a metallic seat and can reduce
the magnitude of impact stresses acting on the sealing surfaces of
the poppets. The valve seat cushioning plate 68 can also be
constructed of a metallic material, such as steel or stainless
steel, which can minimize the wear rate of the conical seats in the
cushioning plate. In either case, a cushioning element can be
positioned between the cushioning plate 68 and the valve seat 11.
As discussed in more detail below, the cushioning element between
the valve seat 11 and the cushioning plate 68 can be a cushioning
pad 73 (see FIGS. 16A and 21), an O-ring 107 (see FIGS. 22A and
22B), a spring 115 (see FIGS. 23A and 23B), or a combination of
thereof.
[0125] Referring to FIG. 22A, an O-ring 107 is shown positioned in
a groove 113 in the valve seat 11. When the cylinder pressure
acting through ports 19 exceeds the control pressure acting through
control pressure ports 71, so that the stem ends of the poppets 13
are held against the valve guard 65, the valve seat cushioning
plate 68 is lifted by the O-ring 107 off of the valve seat 11 by a
distance 114, such that the top surface of the valve seat
cushioning plate 68 is in contact with the valve guard 65.
Referring to FIG. 22B, when the control pressure 46 (arrows)
exceeds the cylinder pressure, the head ends of the poppets 13 are
forced into contact with the conical seats 108 of the valve seat
cushioning plate 68, and the aggregate force applied by the poppets
13 compresses the O-ring 107 and forces the bottom surface of the
valve seat cushioning plate 68 into contact with the valve seat 11.
The resilience and damping provided by the compression of the
O-ring 107 within the groove 113 increases the deceleration time of
the poppets 13 as they come into contact with the conical seats 108
in the valve seat cushioning plate 68, such that the impact stress
on the poppet sealing surfaces is reduced.
[0126] The O-ring cushioning element 107 preferably has a
cross-sectional diameter of at least 0.103 in., more preferably a
cross-sectional diameter of at least 0.201 in., and most preferably
a cross-sectional diameter of 0.275 in., and can be constructed
from an elastomeric material such as fluorocarbon, ethylene
propylene, nitrile, polyester or polyether urethane,
tetrafluoroethylene-propylene or other similar compounds. The
O-ring material can be selected based on the operating conditions
and compatibility with the gas composition and any contaminants,
including lubricants, in the gas stream. The Shore hardness of the
O-ring is preferably at least 85 durometer or higher. The
compressed distance 114 is determined such that each cycle of the
poppets compresses and relaxes the O-ring 107 by a minimum of 1% to
2% of the cross-sectional diameter, preferably a minimum of 3% to
4%, and most preferably a minimum of 5% to 6% of the
cross-sectional diameter.
[0127] Referring to FIG. 21, the cushioning element can also
include a cushioning pad 73 that is very lightly clamped by the
valve guard 65 between the valve seat cushioning plate 68 and the
valve seat 11. The O-ring 107 in this embodiment mainly serves as a
pressure seal that provides minimal cushioning. The damping pad is
constructed of virgin PTFE material having a Shore hardness of
about 50 or a lightly-filled PTFE material having a Shore hardness
of about 60 that is oriented between the valve seat cushioning
plate 68 and the valve seat 11 and aligned by two or more precision
guide sleeves 110, each of which is retained by a cap screw 109 and
sealed by an O-ring 111, such that ports 112 in the cushioning pad
73 align with the ports 108 in the valve seat cushioning plate 68
and the ports 19 in the valve seat 11.
[0128] The cushioning pad 73 thickness preferably ranges from a
minimum of 0.030 in. to a maximum of 0.100 in., and more preferably
between a minimum thickness of 0.060 in. and a maximum thickness of
0.065 in. When the cylinder pressure exceeds the control pressure,
the stem ends of the poppets 74 are held against the sealing
surface of a control chamber spacer plate 64 (which can be
associated with or part of the valve guard 65), and the cushioning
pad 73 is very lightly clamped by the valve guard 65 acting on the
valve seat cushioning plate 68. When the control pressure exceeds
the cylinder pressure and the poppets change direction and move
toward the valve seat cushioning plate 68, the head ends of the
poppets 74 are forced against the valve seat cushioning plate 68,
which creates a compressive force on the cushioning pad 73. The
cushioning pad 73, being of a material that is much softer and more
resilient than the valve seat cushioning plate 68, provides a
cushioning and damping effect that increases the deceleration time
of the poppets as they come into contact with the individual
conical seats in the valve seat cushioning plate 68, such that the
impact stress on the poppet sealing surfaces is reduced.
[0129] Referring to FIGS. 23A and 23B, the cushioning element can
also include a spring 115 positioned in a groove 116 in the valve
seat 11. Looking at FIG. 23A, when the cylinder pressure acting
through ports 19 exceeds the control pressure acting through
control pressure ports 71, so that the stem ends of the poppets 13
are held against the valve guard 65, the spring 115 lifts the valve
seat cushioning plate 68 off the valve seat 11 by a distance 114
such that the valve seat cushioning plate 68 is in contact with the
valve guard 23. Referring to FIG. 23B, when the control pressure 46
(arrows) exceeds the cylinder pressure, the head ends of the
poppets 13 are forced into contact with the conical seats 108 of
the valve seat cushioning plate 68, and the aggregate force applied
by the poppets 13 compresses the spring 115 and forces the valve
seat cushioning plate 68 into contact with the valve seat 11. The
resilience and damping of the spring 115 increases the deceleration
time of the poppets 13 as they come into contact with the conical
seats 108 in the valve seat cushioning plate 68, such that the
impact stress on the poppet sealing surfaces is reduced. The spring
can be a single wafer spring, or a single spiral wound spring,
multiple coil springs acting in parallel, multiple Bellville
washers acting in parallel or any other type of spring or
combination of springs as is known in the art that is sufficiently
compact and provides the required force and spring rate.
[0130] Presented here in FIGS. 24A through 24D is representative
test data that demonstrates the successful high-speed operation of
a variable clearance system that incorporates the improvements of
the present invention including the pressure breaker valve poppet
with extended sealing guide stem (item 74 in FIGS. 16A and 16B) and
the passive valve seat cushioning plate (element 68 in FIGS. 22A
and 22B). The tested compressor cylinder was a double-acting
configuration with a 6.5 in, bore diameter and a 3.0 in. stroke.
The cylinder end closest to the compressor crankshaft (referred to
as the crank end or "CE") operated at full capacity with no
capacity control applied throughout the testing. The end farthest
from the compressor crankshaft (referred to as the head end or
"HE") had a 75.4 in.sup.3 clearance pocket that was connected
utilizing a variable clearance system that incorporated the
improvements and above-described features of the present invention.
This enabled, when operating at higher speeds than was possible
with the prior art, the capacity of the cylinder to be varied by
changing the control pressure Pc within the variable clearance
system.
[0131] FIG. 24A contains plots of measured capacity and the
measured power as percentages of full capacity and full power,
respectively, versus the variable clearance system control pressure
Pc for the previously described double-acting compressor cylinder
operating on nitrogen gas at 1200 rpm with a 200 psia suction
pressure, 90.degree. F. suction temperature and 440 psia discharge
pressure. The control pressure was varied from 205.9 to 433.9 psia
in this data set. The data demonstrates that as the control
pressure was progressively reduced, the compressor cylinder
capacity and the required compression power were reduced as
intended. At this operating condition, reducing the control
pressure reduced the capacity progressively to a minimum of 58.6%
of the combined full capacity of the head end and the crank end of
the cylinder. The required power was similarly reduced
progressively to 66% of the combined full power required for head
end and crank end of the cylinder. The percentage reduction of
power is less than the percentage reduction of capacity because of
parasitic losses generated by the pressure drop of the rapidly
reversing flow through the valve in the variable clearance system,
increasing as the control pressure was reduced. It is noted that
the test results shown in FIG. 24A are for an unloader valve
assembly that did not include the optimized flow path design
relationships described above, which were added after testing to
further reduce the parasitic losses.
[0132] FIG. 24B contains the actual pressure vs. compressor
cylinder volume plots that resulted from setting the control
pressure at 433.9 psia, which is the highest capacity presented in
the data from FIG. 24A. Plot (1) is the actual PV of the CE
cylinder, which is essentially the same as the PV of the HE
cylinder when operating at full capacity without the effects of the
variable clearance system. Plot (2) is the actual PV of the HE
cylinder with the control pressure set at 433.9 psia in the
variable clearance system. Plot (3) is the control pressure as
measured in the control line external to the variable clearance
system. The control pressure measurement location resulted in a
phase lag between the indicated control pressure and the actual
control pressure, however it provides a relative verification of
the control pressure level for reference. Plot (4) is the pressure
inside the clearance pocket volume of the variable clearance
system. It shows that the pressure inside the clearance pocket
decreases with the HE cylinder pressure until the HE cylinder
pressure is slightly less than the control pressure, at which point
the poppets are held closed by the control pressure; and then it is
relatively constant until the HE cylinder pressure slightly exceeds
the control pressure, at which point the poppets open again. The
similar widths of plots (1) and (2) show that the HE is only
slightly unloaded, in this case approximately 8% less than full
capacity of the HE and approximately 4% less than the full combined
HE and CE cylinder capacity.
[0133] FIG. 24C contains similar plots to FIG. B at the same
operating condition that resulted from setting the control pressure
at 205.9 psia, which is the lowest capacity presented in the data
from FIG. 24A. The much narrower plot (2) compared to plot (1)
shows that the HE is significantly unloaded, in this case
approximately 82% less than full capacity of the HE and
approximately 41% less than the full combined HE and CE full
capacity.
[0134] FIG. 24D contains the actual pressure vs. compressor
cylinder volume plots that resulted from setting the control
pressure at 103.9 psia for the previously described double-acting
compressor cylinder and variable clearance system operating on
nitrogen gas at 1400 rpm with a 101.1 psia suction pressure,
90.degree. F. suction temperature and 219.0 psia discharge
pressure. This data is representative evidence that the variable
clearance system, with incorporation of the present invention,
functions reliably at much higher operating speeds (1400 rpm) than
were possible with the prior art variable clearance system
(typically <750 rpm).
[0135] Various poppet and valve seat designs have been described
above for reducing or eliminating poppet leakage and high impact
stresses on the sealing surfaces of the poppets. As noted, the
poppet and valve seat designs are preferably intended for
application with unloader valve assemblies used with reciprocating
compressors operating at 1000 rpm and higher; however, the novel
features disclosed herein can also be applied to improve the
performance and reliability on compressors operating below 1000
rpm. Also, while the present invention is specifically described
for use in unloader valve assemblies at the head or outer end of a
reciprocating compressor cylinder, it is envisioned that the poppet
and valve seat features disclosed herein may also be applied on
suction or discharge valve pockets, or on any pockets that
communicate with the cylinder internal volume on either the head
(outer) end or the frame (inner) end of the cylinder. While the
present invention has been illustrated by the description of
embodiments and examples thereof, it is not intended to restrict or
in any way limit the scope of the appended claims to such detail.
Additional advantages and modifications will be readily apparent to
those skilled in the art. Accordingly, departures may be made from
such details without departing from the scope of the invention.
* * * * *